The present invention relates to balance mechanisms for rotating machinery, particularly balance shafts for multicylinder internal combustion engines which exhibit shaking forces and/or rotating imbalance couples.
Balance shafts are commonly used to reduce or cancel shaking forces and/or vibrations which result from residual imbalances inherent in the design architecture of machinery with rotating parts or mechanisms, such as motors. These balance shafts are sometimes called xe2x80x9ccounterbalancexe2x80x9d shafts.
Balance shafts are particularly valuable when operator or passenger comfort and freedom from noise and vibration-related fatigue or distraction are desired, as in the case of motor vehicles such as automobiles, motorcycles, and the like. It is also advantageous to minimize vibration from the standpoint of equipment reliability. Where vibrations are reduced, the size, mass and/or complexity of the mounting structures can often also be reliably reduced, thus potentially reducing cost.
With multicylinder motor vehicle engines, the inline four-cylinder engines and 90-degree V-6 engine configurations are favored in automotive use today due to their space efficiency and cost. Both of these engine architectures benefit from balance shafts, although for different reasons and vibratory characteristics, and thus requiring distinctly different balance shaft arrangements.
Balance shafts for inline four-cylinder engines typically are paired to rotate in opposite directions at twice the engine speed. The two balance shafts are timed to cancel each other""s lateral shaking forces while opposing the vertical secondary shaking forces that are typical with this type of engine. Each shaft produces a single, or xe2x80x9cstatic,xe2x80x9d rotating unbalance force, which taken together with its mating shaft""s rotating unbalance force, produces a resultant vertical shaking force which most effectively is located centrally among the bank of cylinders. These static unbalance type shafts are shown, for example, in U.S. Pat. No. 4,819,505.
Other engines, such as 90-degree V-6 engines (i.e., six-cylinder engine with two banks of three cylinders spaced 90-degrees apart), produce resultant imbalance forces in the form of a crankshaft-speed rotating couple. These engines benefit from a single balance shaft with two balance xe2x80x9cweightsxe2x80x9d, or masses, on opposite sides of its axis of rotation, but spaced apart axially so as to have a dynamic imbalance providing a rotating couple. The couple produced by the balance shaft is designed to oppose or cancel that of the engine when the shaft is rotated at crankshaft speed and in the opposite direction to the crankshaft. The location of this xe2x80x9crotating couplexe2x80x9d-type shaft relative to the engine is not critical so long as its axis of rotation parallels that of the crankshaft, since the output of the balance shaft is a pure couple or torque on the crankcase.
Balance shafts of both types frequently incorporate an elongated support member, or shaft, which provides a structural connection between the balance weights, in the case of rotating couple-type shafts, or between the centrally located balance weight(s) and a driving member, in the case of the static unbalance-type shaft. The elongated support member is typically subjected to both torsion and bending loads, and thus must be substantial enough to fulfill structural requirements. Since the mass of the elongated support member is largely xe2x80x9cdead weightxe2x80x9d and has little, if any, contribution to unbalance, its mass can be reduced in applications where overall mass is a factor in product cost and/or operating efficiency. These elongated support members or shafts typically have a circular cross-section. This circular section represents a structurally inefficient distribution of material that causes the components and their support structures to be more massive and often more costly than necessary.
The room or space for placement of balance shafts in the engine is typically small or limited. Balance shafts usually are constrained to operate within specified radii, whether to clear mating parts or to enable installation. Thus, efficient material usage typically motivates a balance weight cross-sectional shape that is, except for elongated support member intersection areas, xe2x80x9ccircular segmentxe2x80x9d in shape, i.e. the area between a radius and a chord. The radius of such a shape represents the clearance boundary beyond which the balance shaft cannot extend without risk of unwanted contact. The chord represents a locus of constant contribution to unbalance within the section, placing elements of mass equidistant from the axis of rotation, with regard to the ability of the mass element to generate centrifugal force in a particular direction, i.e., when viewed from a direction normal to the desired direction of unbalance force.
Typically, the xe2x80x9ccircular segmentxe2x80x9d shape of the balance weights are constant along their lengths. This enables easy calculation of their unbalance value from a design standpoint. However, this shape also results in inefficient distribution of material in the case of shafts with balance weights which create a rotating couple, or dynamic imbalance, thus causing components and their support structures to be more massive and thus also often more costly than necessary.
Space constraints sometimes preclude the placement, within the inline four-cylinder type engine and in conjunction with appropriate structural support, of balance weights in a manner that results in the resultant vertical shaking force being located centrally among the bank of cylinders as desired. In this situation, an unwanted pitching couple is created as a result of the axial distance between the engine""s vertical shaking force and the balance shafts"" resultant vertical shaking force, unless additional balance weights can be added to create rotating couple, or dynamic, unbalance within each shaft that will act to cancel this pitching couple. Such dynamic balance, when added to a static unbalance-type shaft can be seen to effectively relocate the plane of static unbalance to the new axial location where the sum of the moments of unbalance, or dynamic unbalance, within the shaft itself is zero. Any such combination of static and dynamic unbalance within a shaft can thus be characterized by an amount of pure static unbalance at an effective location or plane hereafter referred to as its xe2x80x9cEffective Plane of Static Unbalancexe2x80x9d, or xe2x80x9cEPSUBxe2x80x9d, about which the sum of moments of unbalance is zero.
The ideal application of balance shafts to inline four cylinder engines will locate the shafts, EPSUB at the axial center of the four cylinders, such that no pitching couple is created by an offset between the engine""s shaking force and the balance shafts"" shaking force, or in other words the sum of shaking force moments about the engine""s axial center is zero. Where space constraints prevent this ideal full cancellation, the resulting residual shaking force may be located optimally by similar EPSUB methodology so as to most appropriately distribute the residual shaking force among engine mounts using appropriate noise, vibration and harshness minimization criteria.
Manufacturing cost consideration often force design compromises between ideal bearing configurations and ideal balance weight configurations. For example, it is common to use a larger than optimum (for friction losses, heat generation, etc.) bearing journal diameter in conjunction with a balance weight clearance boundary radius that is smaller than optimum (for unbalance creation without undue material usage) to enable axial installation (or xe2x80x9cend loadingxe2x80x9d) of the balance weight through the bearing bore, rather than incur the manufacturing complexity and cost associated with the split housing type bearings required to place an ideal configuration bearing in the midst of two larger radius balance weights that are symmetrically arrayed about the engine""s center bulkhead.
The common method for providing for bearing journal diameter(s) smaller than balance weight radius without requiring split housing type bearings, namely fastening weights to a shaft after inserting the shaft through its bearing(s), is also complex, and thus also costly to manufacture, as well as being heavier than necessary.
There exists, therefore, potential for improvement in reducing manufacturing cost and solving space constraint problems, while managing the issues of drive system noise, bearing reliability, bearing drag, and overall weight in a manner that maximizes product value to the customer in the use of static unbalance balance shafts.
It is the object of the present invention to provide improved balance shafts for rotating machinery such as motor vehicle engines by enabling balance shaft design configurations which:
1.) result in lighter weight, and thus also potentially lower cost, by means of improved utilization of material in the elongated support member areas of the component for given load conditions;
2.) are stronger, having greater factor of safety for a given material usage, by means of improved utilization of material in the elongated support member areas of the component;
3.) contribute to increased bearing life due to the reduced bearing journal tilt angles that result from increased stiffness (resistance to bending under centrifugal loads) for a given material usage, by means of improved utilization of material in the elongated support member areas of the component;
4.) exhibit increased stiffness (resistance to bending under centrifugal loads) by means of improved utilization of material in the elongated support member areas of the component, with the associated benefit of reduced bearing journal tilt and thus potentially increased operating efficiency by means of smaller, and thus lower drag, bearing sizes;
5.) result in lighter weight and thus also potentially lower cost by means of improved utilization of material in the balance weight areas of shafts which create a rotating couple;
6.) reduce parasitic power loss by means of reduced xe2x80x9cwindagexe2x80x9d, or drag from air resistance, due to the reduced xe2x80x9cfrontal areaxe2x80x9d and bluntness of smaller, more efficiently shaped balance weights which create a rotating couple;
7.) reduce gear size and cost requirements as needed to achieve quiet operation through elimination of need to counteract the effects of unwanted operating deflections, which also influence bearing size requirements and thus cost;
8.) minimize bearing drag, which increases as the cube of bearing journal diameter, which in turn is driven by considerations of journal tilt under unbalance loads, with tilt magnitude being a function of shaft stiffness and the distribution of unbalance-creating material; and/or
9.) reduce manufacturing cost while meeting space constraints without inappropriate penalties to functional priorities of assuring bearing reliability, minimizing drive system noise, minimizing frictional losses, and minimizing overall weight.
The present invention enables the above object to be achieved by providing design methods and structures which result in improved balance shaft configurations, having reduced friction, and potentially reduced weight and/or manufacturing cost, with improved operating shapes under centrifugal bending loads, with potential attendant benefits of improved bearing reliability. Reduced weight can allow for subsequent weight reductions in associated support structures of the engine or vehicle.
In accordance with one embodiment of the present invention, the cross-sectional shape of the elongated support member or shaft, hereafter referred to as the xe2x80x9cconnector portionxe2x80x9d, between the balance weight(s) and the driving means of the static unbalance-type balance shaft, is formed in an optimized manner to minimize material usage while maintaining required bending stiffness, torsional stiffness, and safe levels of mechanical stress. The cross-section of the connector portion is shaped substantially like an xe2x80x9cI-beamxe2x80x9d with recessed or concave portions. This improves the ratio of section modulus to mass in the direction of the centrifugal loads, which in turn reduces the peak stress for a given material usage. Optimization of the connector portion may involve tapering, such that the xe2x80x9cI-beamxe2x80x9d varies in section along its length to address the variation in bending moment along its length.
As to another embodiment of the present invention, namely balance shafts with balance weights that create a rotating couple, one of the surfaces on each of the balance weights of the shaft is preferably shaped as a hyperbolic curve or an approximation thereof. The hyperbolic curve represents the locus of constant contribution to the unbalance couple produced by the shaft. There is a unique and preferred hyperbolic curve for each combination of unbalance value and balance weight clearance boundary conditions.
The cross-sectional shape of the connector portion between the balance weights of the static with rotating couple-type balance shaft is also formed in an optimized manner to minimize the material usage. The cross-section of the connector section is shaped substantially like an xe2x80x9cI-beamxe2x80x9d with recessed or concave portions. This improves the ratio of section modulus to mass in the direction of the centrifugal loads, which in turn reduces the peak stress for a given material usage.
Still further embodiments of the present invention provide improved static unbalance-type balance shafts, some with counterweights which overhang one of the bearing journals, and some with a combination of static unbalance and rotating couple-type configurations. The static unbalance-type balance shafts reduce material volume while improving operating deflection shape for the benefit of bearing reliability and/or gear noise and/or gear size and cost requirements necessary for quiet high speed operation by means of elongating balance weights from their typical rectangular side view proportions, in conjunction with longitudinal direction tapering of their (sectionally substantially chordal) inner surfaces. The elongation of the more effective (toward unbalance creation) outer portion (near clearance boundary radius, when viewed normal to direction of unbalance and axis of rotation) of the balance weights in conjunction with longitudinal tapering of the inner surfaces to maintain equivalent unbalance value serves to reduce mass, while increasing bending stiffness, in the case of the balance weight(s) between journals, and while potentially reducing bearing journal tilt under high speed unbalance loads of the xe2x80x9coutriggerxe2x80x9d bearing and its adjacent drive means, in the case of the overhung balance weight. To the extent that the moment of unbalance, about the length centerline of the principal bearing, of the overhung balance weight exceeds that of the balance weight between support bearings, the principal bearing is used as a fulcrum to offset the bending deflection of the shaft between bearings, to the potential straightening, under high speed operating loads, of the outrigger journal and its adjacent drive means, which can be of critical importance in the maintenance of the theoretical, or undeflected, helical contact ratio of drive gears and/or coupling gears as required for quiet operation.
If helical gearsets are not operated in high states of parallelism, i.e., freedom from errors due to manufacturing tolerances and operating deflections, the (theoretically) line contact upon which helical, and thus total, contact ratios of gearsets depend is reduced to (theoretically) point contact at the edges of the gears. To the extent gear faces are crowned to accommodate non-parallelism, the (theoretically) line contact is reduced to (theoretically) point contact anyway, to the effective loss of helical contact ratio and thus total contact ratio.
Journal tilt magnitudes are also a design consideration in the optimization of support bearings, with plain, or journal-type bearings especially susceptible to edge loading as a principal cause of seizure failures. Symmetry of balance weight distribution about a principal load carrying journal has been the traditional approach in the effort to minimize journal tilt, but the reality of high speed operating deflections as predicted by computer simulation such as Finite Element Analysis (FEA) often reveals that this symmetry approach fails to achieve the intended results because of failure to account for the effects of shaft or connector portion stiffness between bearings.
A preferred embodiment of the present invention provides for the manufacturing simplicity and cost benefits of axial assembly of one-piece, two-journal balance shafts to unsplit housing bearings, along with the friction loss benefits of bearings which can be of ideal size and configuration. Challenging space constraints are potentially also met with fewer compromises to clearance boundary radius, by locating static unbalance-type shafts which incorporate dynamic unbalance (to effect the appropriate EPSUB location at, or near, the central bulkhead of an inline four cylinder engine), in either the front or rear half of the engine. Low mass technology disclosed herein and in original U.S. application Ser. No. 08/677,085 can be utilized to minimize the shaft weight despite inclusion of the added dynamic unbalance which eliminates the necessity of split housing type bearings by eliminating the need to distribute unbalance mass on both sides of the principal (engine center bulkhead area) bearing journal.
Other benefits, features and advantages of the present invention will become apparent from the following written description of the invention, when taken in accordance with the appended claims and accompanying drawings.